Cam driven compressor



y 22, 1969 w. GRAPER 3,456,874

CAM DRIVEN COMPRESSOR Filed Aug. 1, 1967 5 Sheets-Sheet 1 I NV ENTOR. WILLIAM GRAPER ATTORNEY July 22, 1969 w. GRAPER CAM DRIVEN COMPRESSOR 5 Sheets-Sheet 2 Filed 1, 1967 I NV ENTOR.

WILLIAM GRAPER ATTORNEY July 22, 1969 w. GRAPER 3,456,874

CAM DRIVEN COMPRESSOR Filed Aug. 1, 1967 5 Sheets-Sheet 8 ATTORNEY July 22, 1969 w. GRAPER 3,456,874

CAM DRIVEN COMPRESSOR Filed Aug. 1, 1967 Y 5 Sheets-Sheet. 4

INVFNTOR WILLIAM GRAPER ATTCRN E Y CAM DRIVEN COMPRESSOR Filed Aug. 1, 1967 I 5 Sheets-Sheet 5 INVENTOIL WILLIAM GRAPER ATTORNEY United States Patent U.S. Cl. 230-185 7 Claims ABSTRACT OF THE DISCLOSURE A cam driven compressor having hydrostatic bearings and dynamic balancing allowing high speed operation and relatively low noise level and with a minimum of bearing friction.

BACKGROUND OF INVENTION Field of invention This invention relates to a novel cam driven compressor generally, and more specifically, to a cam driven compressor having novel bearing lubrication means to permit high speed operation.

Description of the Prior art Compressors utilizing an eccentric cam and sliding block driving arrangement are generally old in the art as shown and illustrated in US. Patent 1,435,224 issued Nov. 14, 1922, to W. Gensecke. Such compressor drives would appear highly desirable because of the simplicity of construction and relatively few moving parts. In fact, the simplicity of construction and relatively few moving parts of a cam driven compressor are highly desirable chracteristics in compressors used in todays applications. However, problems such as inadequate bearing means and inherent unbalance have relegated cam driven compressors to use as low speed fluid pumps and the like. No high speed cam driven compressor has been successfully commercially marketed because of the above noted problems. The problem of providing a satisfactory bearing for high speed operation has been a major obstacle to a successful cam driven compressor. The problems associated with the cam driven compressors of the prior art have been solved by the invention of the present case. A highly improved lubrication system is provided together with a novel lubrication means at the bearing surfaces of the compressor to allow sustained high speed operation of the compressor of the present case. The novel bearing arrangement and lubrication means also serves to allow high speed operation of the compressor of the instant case at relatively low noise levels.

SUMMARY OF INVENTION In a compressor of the type having a piston driven by an eccentric circular cam through a sliding block, the improvement of a hydrostatic lubrication and bearing system between the rotating and sliding surfaces of the compressor drive mechanism allowing sustained high speed operation without undue friction, noise and vibration.

BRIEF DESCRIPTION OF THE DRAWINGS FIGURE 1 is a cross sectional view of a cam driven compressor showing certain elements and features of this invention.

FIGURE 2 is a transverse sectional view of the compressor shown in FIGURE 1 taken along the section line 22.

FIGURE 3 is a partial sectional view of the compressor taken along the section line 33 of FIGURE 1 illustrating further aspects of this invention.

FIGURE 4 is a partial sectional view of the compressor taken along the section line 44 of FIGURE 2 and illustrating still further details and aspects of this invention.

FIGURE 5 is a plan view taken along the lines 55 of FIGURE 2 showing the inlet and exhaust valve and port arrangement of the compressor of the instant invention.

FIGURE 6 is a plan view taken along the section line 6-6 of FIGURE 2 illustrating further details of the exhaust valve.

FIGURE 7 is a plan view taken along the section line 7-7 of FIGURE 2 illustrating in greater detail the inlet valve of the compressor of the present invention.

FIGURE 8 is a plan view taken along the line 88 of FIGURE 2 and illustrating further aspects of the inlet and exhaust valves of the present invention.

FIGURE 9 is a plan view of the cam block taken along the line 9 of FIGURE 2 and illustrating in detail an important feature of the present invention.

FIGURE 10 is a lubrication pumping mechanism used in the present invention.

FIGURE 11 is an overall view of the compressor of the present invention and illustrating further aspects of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENT Shown in FIGURES 1-11 is a typical compressor providing an environmental background in which the present invention is operable. Such compressor is designated generally by reference numeral 20.

An important component of the compressor is a generally rectangular primary housing 22 having a cylindrical bore 24 therethrough. The primary housing 22 is closed at each of the axial ends by first and second head members 26 and 28 and first and second end caps 30 and 32. Located adjacent either side of the compressor housing 22 are first and second bearing support members 34 and 36.

Attached to the first bearing support member 34 is a cover plate 38.

The various elements of the compressor housing are associated in the following manner to perform the following described functions. "I he first and second head members 26 and 28 are located against the primary housing 22 and sandwiched by the first and second end caps 30 and 32. The head members 26 and 28 and the end caps 3t) and 32 are secured to the primary housing 22 by a plurality of removable fasteners 40. The first and second head members 26 and 28 cooperate with the primary housing 22 and a double acting piston member 42 to define a pair of air compression chambers 44 and 46 within the cylindrical bore 24 of the primary housing 22. The first and second end caps 30 and 32 are provided with a plurality of inlet ports 48 and exhaust ports 50. Flow through the inlet ports 48 is controlled by inlet valves 52.

The inlet valves 52 are generally strip shaped spring material and extend across the cylindrical bore 24 of the compressor housing 22. The inlet valves 52 are pinned at one end portion 53 to the head members 26 and 28, by inlet valve dowels 54. Another end portion 56 of the inlet valves 52 is located within a recess 58 in the housing 22. The recess 58 serves to limit the travel of the intake valve 52 and further serves as a support for the spring action of the intake valves 52.

Also mounted on the head members 26 and 28 are exhaust valves 60 which cover the exhaust ports 50. As shown in FIGURES 5, 6 and 8, the exhaust valves 60 are located over the exhaust ports and are mounted on the head members 26 and 28 by dowel pins 62 located intermediate the ends of the exhaust valve 60. Also mounted on the head members 26 and 28 by means of the dowel pin 62 are exhaust valve backup members 64 which limit the travel of the exhaust valve 60 in an opening direction.

The end caps 30 and 32 have a transverse web 66 which cooperates with the head members 26 and 28 to define an inlet passage 68 and an exhaust passage 70.

The inlet passage 68 formed by the end caps 30 and 32 is connected wit-h a passageway 72 in the primary housing 22 and with the atmosphere by means of inlet port 74.

The inlet port 74 and exhaust port 78 are surrounded by generally oval land portions 80 and 82 for the connection of appropriate fluid supply and delivery fittings, not shown. Because of the continuous nature of the inlet and exhaust passages, it is possible, as shown in FIGURE 11, that the inlet and exhaust lands 80 and 82 be located in various positions on the compressor such as at 80', 82. The optional location of the lands 80 and 82 allows a considerable amount of freedom in the application of the compressor of the present invention.

Located within the bore 24 of the primary housing 22 is the compressing piston 42. The piston 42 is double acting and is driven in a reciprocating manner through a cam block 84. The cam block 84 is driven by an eccentrically mounted circular cam 86 mounted on a cam shaft 88 which is driven by any suitable rotating power source such as an internal combustion engine.

Thecam shaft 88 extends through the primary compressor housing 22 and into the second bearing support member 36 where the cam shaft 88 is supported in a bearing 92. Adjacent a journal 90 on the cam shaft 88 is a counterweight 94. Also located within the primary cornpressor housing 22 is a second counterweight 96 which is removably secured to the cam shaft 88 to allow easy assembly of the cam shaft with the remainder of the compressor. The second counterweight 96 is removably secured to the cam shaft 88 by means of fasteners 98. A ball bearing 100 is mounted in the first bearing support member 34 and supports the cam shaft 88 for rotation.

Also located on the cam shaft 88 is a bushing 102 and a seal 104. The bushing 102 and seal 104 are located within the cover plate 38 and prevent the leakage of lubrieating fluid and the entry of dirt to the bearing 100.

Connected to the cam shaft 88 and supported by the second bearing support member 36 is a pump assembly 106 for supplying lubricating fluid to the various bearing elements of the compressor. The pump assembly 106 comprises a rotor member 108 mounted on a snout portion 110 of the cam shaft 88. Located within an eccentric bore 112 of the second bearing support member 36 is a freely rotatable driven member 114. The pump assembly with the above identified elements comprises a pumping arrangement commonly known in the art as a Gerotor pump. Also provided in the second bearing support member 36 is a lubrication intake passage 116 and an exhaust passage 118. The inlet passage 116 is connected to a sump 120 by a conduit 122. The exhaust passage 118 is connected by internal ducting to a port 125 located adjacent the snout portion 110 of the cam shaft 88.

The cam shaft 88 contains a bore 124 for the delivery of lubricating oil to the ball bearing 100 and to the slid- I ing block 84 by means of a secondary bore 126.

The sliding block 84 is provided with oil delivery passages 128 for the delivery of oil from the lubricating pump assembly 106 to the bearing surface between the cam block 84 and the piston 42. The cam block 84 as shown in FIGURE 9 is provided with lubricating grooves 130 on the bearing surfaces 132 thereof.

Having thus fixed the location of the various elements of the compressor, reference is now made to FIGURES 1 through 11 for a more specific discussion and description of the various elements and their functions in the compressor.

As shown in FIGURE 1, the cam shaft 88 is adapted to be rotatably driven within the compressor housing 22 by an external power source not shown. As can be seen from FIGURE 1, the cam shaft 88 will rotate about an axis defined by the ball bearing 100 mounted in the cover plate 34 and the journal bearing 92 located in the second cover plate 36. Located intermediate the ball bearing 100 and the journal bearing 92 and as indicated above, is a circular cam element 86 which is mounted eccentric of the axis of rotation of the cam shaft 88. As can be better seen in FIGURE 2, located concentric with the circular cam element 86 is a sliding cam block 84 adapted to be driven by the cam element 86. It can be seen that as the cam shaft 88 is rotated about the above described axis of rotation the cam element 86 will drive the cam block 84 in a circular motion with relative movement therebetween. The circular motion of the cam block 84 imparted thereto by the eccentric cam member 86 is used to drive the piston 42 in a reciprocating manner. In providing for the driving of the piston element 42 by the cam block 84 there must, of course, be provided a bearing surface 132 between the cam block 84 and the piston 42 in order to allow relative motion t-herebetween. The bearing surface 132 is needed since a portion of the circular motion of the cam block 84 is lateral to the axis of the reciprocating piston 42.

Having thus described the translation of the rotary motion of the cam shaft 88 to reciprocating motion in the piston 42, the air flow into and from the compressor chambers 44 and 46 may be described thusly: rotation of the cam shaft 88 may, of course, be in either direction, but for purposes of illustration and description, rotation of the cam shaft 88 in the direction of the arrow in FIG- URE 2 has been assumed. Rotation of the cam shaft 88 from the position shown in FIGURE 2 will cause the cam block 84 to move upwardly and to the right and moving the piston 42 therewith. When the cam shaft 88 has been rotated through an angle of the cam block 84 and the piston 42 will be at their maximum upward displacement. The maximum displacement of the piston 42 is governed by the dimension of eccentricity 2 shown in FIGURE 2 of the drawings. The maximum displacement of the piston 42 is twice the eccentricity e of the cam element 86. Further rotation of the shaft 88 from the vertical position of the axis of eccentricity of the cam element 86 will cause the piston element 42 to begin its downward movement. Since the inlet 'valve 52 and the exhaust valve 60 normally maintain the compression chambers 44 and 46 in an air tight condition, the downward movement of the piston 42 will cause a partial vacuum to exist in the compression chamber 44. Since the inlet ports 48 are open to the atmosphere through the passages 68 and 72 through the inlet port 74, the inlet valves 52 mounted on the head member 26 will be forced open by the pressure differential existing in the passageway 68 and the air compression chamber 44. This pressure differential will continue to exist until the cam shaft has rotated 270 from the position shown in FIGURE 2. Therefore, air will continue to be forced into the compression chamber 44 by the pressure differential existing and the compression chamber 44 will be thus charged with air at slightly below atmospheric pressure. Continued rotation of the cam shaft 88 from a position 270 away from the position shown in FIGURE 2 of the drawing, will cause the piston 44 to start in an upward direction. During upward movement of the piston 42 the air contained in the compression chamber 44 will be compressed. The degree of compression of the air in the chamber 44 will depend upon the force with which the exhaust valve 60 is held against the head member 26. In other words, the air in the compression chamber 44 as the piston 42 is moving upwardly is statically compressed until the area of the exhaust port 50 multiplied by the pressure in the compression chamber 44 exceeds the force necessary to lift the exhaust valve 60 from the head member 26.

Since the piston member 42 is double acting it is readily apparent that air is being forced into and compressed in the compression chamber 46 in a manner identical to that described immediately above. It is to be further understood, of course, that the intake and compression of the air in chamber 46 is 180 out of phase with the corresponding action of the piston on the air in the compression chamber 44. Upon reaching the necessary pressure to lift the exhaust valve 60 from the head member 26, the compressed air is then forced into the passages 70 and 76 and out through the exhaust port 78 in the primary housing 22 where it is then used for the purposes desired.

It can be readily seen that during rotation of the cam shaft 88 there exists relative movement between the eccentric cam 86, the sliding block 84, and the piston element 42. It is the existence of this relative movement between the various parts that has lead to the invention of the present case. In order to minimize the friction existing between the relatively moving surfaces and to reduce the accompanying wear attributable thereto, a novel bearing and lubrication device is provided.

As is best illustrated in FIGURE 3 of the drawing, a sump 120 is provided in the primary compressor housing 22 to receive a quantity of lubricating fluid. Extending from the second bearing support member 36 to the sump 120 is a conduit member 122 which is provided to allow the flow of lubricating fluid from the sump to the inlet passage 116 in the bearing support member 36. The lubrication inlet passage 116 is in communication with the inlet side of a pumping mechanism 106 shown in FIG- URE of the drawing.

The pumping mechanism 106 comprises a rotor element 108 which is non-rotatably secured to the snout portion 110 of the cam shaft 88 and a rotatable element 114 I which is eccentrically located in the bearing support member 36 with respect to the axis of rotation of the cam shaft 88. The operation of a pump mechanism of this type is generally understood to those skilled in the art. The outer pumping element 114 is provided with more teeth than the rotor element 108. As the rotor element 108 is caused to rotate by the cam shaft 88, the outer pumping element 114 rotates in the same direction as the rotor element 108 at a speed related to the difference in the number of teeth between the outer pumping element 114 and the rotor element 108. The relative rotation of the rotor element 108 and the outer pumping element 114 provides a pumping action.

As shown in FIGURES l and 3 the pumping action of the pumping mechanism 106 is used to supply pressurized lubricating fluid to the chamber 125 in the bearing support member 36. From the' chamber 125 the lubricating fluid is forced through the bores 124, 126 and 127 in the cam shaft 88 for the purposes of lubricating the ball bearing 100 and the bearing surface existing between the eccentric cam 86 and the sliding block 84. The journal 90 and the bearing 92 in the bearing support member 36 are lubricated by leakage of the pump assembly 106. As the lubrication of the bearing surface between the cam element 86 and the sliding block 84 and between the sliding block 84 and the piston 42 constitutes a major portion of this invention, the precise nature of the method of lubrication existing between these various elements will be described here in detail.

As the lubricating fluid is forced through the bore 124 in the cam shaft 88 by the pumping means 106 the fluid flows radially outwardly from the bore 124 through a secondary lubrication passage 126 in the eccentric cam member 86. From the bore 126 in the cam shaft 88 the lubricating fluid flows in to the circumferential channel 136 in the sliding block member 84. This channel of pressurized lubricating fluid between the cam element 86 and the sliding block 84 provides for pressurized lubrication between the cam element 86 and the sliding block 84 along the bearing surface 138. The lubricating channel 136 is also in communication with a pair of lubricating bores 128 in the sliding block member 84. The bores 128 communicate with a pair of diagonal grooves 130 on the bearing surfaces 132 of the sliding block 84. Oil under pressure flows from the lubricating channel 136 through the bores 128 and into the diagonal grooves 130 of the sliding block member 84. The pressurized fluid attempts to lift the bearing surface 132 of sliding block from the bearing surfaces 134 of the piston member 42. There then exists between the sliding block member 84 and the bearing surface 134 of the piston member 42, a lubricating fluid film which prevents contact of the sliding block member 84 with the bearing surface 134 of the piston member 42. Since there is no contact existing between the' sliding block 84 and the bearing surface 134 of piston 42 it is readily apparent that there is a substantial reduction in friction existing therebetween. Once the lubricating fluid has served the purpose of lubricating the various bearing elements 100, 92 and the bearing surfaces existing between the cam element 86 and the sliding block 84 and between the sliding block 84 and the bearing surface 134 of the piston member 42, the lubricating fluid returns in a generally random manner to the sump of the primary compressor housing 22.

An important discovery made during this invention was that the sliding block 84 and the piston 42 must be constructed of materials having substantially similar coefficients of expansion in order to prevent the sliding block 84 and the piston 42 from having in the alternative, excessive clearance or insuflicient clearance at normal operating temperatures of the compressor, once an initial satisfactory clearance has been established. Excessive clearance, of course, results in a breakdown of the hydrostatic bearing surface by preventing proper pressurization of the lubricating fluid and further results in excessive noise during operation. Insuflicient clearance likewise results in improper compressor operation by preventing the proper formation of a hydrostatic bearing surface and thereby results in increased friction and wear between the sliding block 84 and the piston 42.

Another important feature of the present invention is the provision of a removable counterweight 96 on the cam shaft 88. The removable counterweight 96 facilitates the easy assembly of the sliding block member 84 and the piston member 42 on the eccentric cam portion 86 of the cam shaft 88. Further, the removable counterweight 96 facilitates the easy assembly of the cam shaft with the remainder of the compressor by allowing assembly of the compressor through the opening covered by bearing support member 34.

Having thus described my invention, it will be immediately apparent to those skilled in the art that a significant contribution to the state of the art of cam driven compressors has been made and that applicant is entitled to protection afforded by the appended claims.

Having thus described my invention, I claim:

1. A compressor for pressurizing fluid comprising:

a housing having a bore therein and defining a plurality of fluid flow passages;

a piston reciprocally mounted in said bore and cooperating with the housing to define a compression chamber;

a sliding block member operatively connected with the piston in sliding, driving relationship therewith, said operative connection including a predetermined flow restrictive clearance between the piston and the block and the sliding block member further having a circular bore therethrough;

a cam shaft assembly mounted on the housing and including a circular cam element eccentrically attached thereto, said cam element being located concentrically with the sliding block member and in driving relationship therewith; and

a pump assembly mounted on the housing and driven by the cam shaft for supplying a quantity of pressurized fluid to the flow restrictive clearance between the piston and the sliding block member and between the sliding block member and the circular cam element,

whereby pressurized fluid supplied to said flow restrictive clearance results in a fluid film support between the piston and the sliding block member.

2. The compressor of claim 1 wherein:

the piston and the sliding block member are constructed of materials having substantially the same coefficient of thermal expansion.

3. The compressor of claim 1 wherein the sliding block member includes a groove on a surface thereof facing the flow restrictive clearance for receiving said pressurized fluid and preventing contact of the sliding block member with the piston.

4. The compressor of claim 3 wherein the sliding block member has grooves extending over a substantial area of the sliding block member surface facing the piston.

5. The compressor of claim 3 wherein:

the piston and sliding block member are constructed of materials having substantially the same coeflicients of thermal expansion.

6. The compressor of claim 4 further comprising:

means for providing a hydrostatic bearing surface between the cam element and the sliding block member.

7. The compressor of claim 6 wherein the means for providing a hydrostatic lubrication bearing surface between the sliding block member and the cam element comprises:

the sliding block member having a circumferential groove in the bore thereof in communication with the source of pressurized lubricating fluid; and wherein the sliding block member and the cam element cooperate to define a flow restrictive clearance therewhereby a fluid load bearing film is maintained between the sliding block member and the cam element.

References Cited UNITED STATES PATENTS 10 between,

Marcus et a1 230- Dunning 230-185 Fourness 230-185 Olcott 230-206 Phelps 230-206 Kropiwnicki 230-206 Chew 230-206 US. Cl. X.R. 

